A loader vehicle, for example, a wheel loader (hereinafter, simply called "the vehicle") has a working machine at the front portion of the vehicle body. The working machine has a boom, with the rear end of the boom being coupled to the front portion of the vehicle body via pivot pins, and a bucket, coupled at the front end of the boom via pivot pins. As shown in FIG. 18, the boom is rotatable around the pin coupling at the rear end of the boom via the operation of the boom cylinder 11, and the bucket is rotatable around the pin coupling at the front end of the boom via operation of the bucket cylinder 12. Specifically, the vehicle has a working machine hydraulic circuit 10. It is normal to include a steering hydraulic circuit 20 together with the hydraulic circuit 10.
Incidentally, the steering circuit, as well as the brake, is a safety element of a vehicle. For this reason, in some vehicles, the working machine hydraulic circuit 10 receives pressurized oil discharged from a working machine hydraulic pump 30 at a flow rate Qw, and the steering hydraulic circuit 20 receives pressurized oil discharged from a steering hydraulic pump 40 at a flow rate Qst, whereby the circuits 10 and 20 are independent from each other. Such a vehicle further has an auxiliary hydraulic pump 50 and a flow dividing valve 60, which is connected to the pumps 40 and 50 and to the circuits 10 and 20 in order to use the engine torque efficiently. The operation of the flow dividing valve 60 is as follows.
At a medium or higher speed of the engine 70, the flow Qst of pressurized oil, discharged from the steering hydraulic pump 40, is sufficient for the steering hydraulic circuit 20. Accordingly, at this time, the flow dividing valve 60 supplies pressurized oil at a flow Qs from the auxiliary hydraulic pump 50 to the working machine hydraulic circuit 10. However, when a steering operation is carried out at a low speed of the engine 70 (at so-called minimum idling engine speed), the oil flow rate Qst is not sufficient for the steering hydraulic circuit 20, and a quick steering operation cannot be conducted. The flow dividing valve 60 then supplies pressurized oil, discharged from the auxiliary pump 50, to the steering hydraulic circuit 20 at the flow rate Qs. Incidentally, oil which has flowed through either of the circuits 10 and 20 returns to a tank 90 by way of a drain circuit 80. Each of the hydraulic pumps 30, 40, and 50 is a fixed displacement type. Some vehicles are equipped with variable displacement type hydraulic pumps, but their variable control does not serve the purpose of the present invention, the details of which will be described below; therefore, in the present invention such a variable displacement type hydraulic pump is considered as a fixed displacement type.
In the vehicle, the operating oil pressure Pw becomes lower during an operation with a low load (for example, when the boom is being raised with a load in the bucket). At this time, the higher the oil flow in the working machine hydraulic circuit 10, the higher the boom ascending speed becomes, and the higher the working speed becomes. On the other hand, during an operation with a high load (for example, during rock excavation by the bucket), the oil pressure of the working machine hydraulic circuit 10 (hereinafter referred to as "operating oil pressure Pw") becomes higher. At this time, the necessary flow of pressurized oil in the working machine hydraulic circuit 10 may be smaller. Then there is an example having an unload circuit 100 to which the oil flow Qs, heading for the working machine hydraulic circuit 10 from the flow dividing valve 60, is drained to the tank 90 when the operating oil pressure Pw exceeds a previously specified oil pressure Pu (hereinafter referred to as "unloaded condition starting pressure Pu"), and is raised to be higher (specifically "Pmax.gtoreq.Pw&gt;Pu"). Here, "Pmax" is the relief oil pressure of the working machine hydraulic circuit 10. The details are as follows.
For example, the unloaded circuit 100 in FIG. 18 has: a check valve 101, which opens in one direction toward the working machine hydraulic circuit 10 from the flow dividing valve 60; an oil passage 102, branching from a portion between the flow dividing valve 60 and the check valve 101 and connecting to the tank 90; and an on-off valve 103, provided in the oil passage 102. The on-off valve 103 has a spring 104, which is initially set with momentum corresponding to the unloaded condition starting pressure Pu. The on-off valve 103 is a two position switching valve which can be switched according to the magnitude of the operating oil pressure Pw by receiving the operating oil pressure Pw from the working machine hydraulic circuit 10 to oppose the momentum given to the spring 104. Specifically, when the operating oil pressure Pw is defined by "Pw.ltoreq.Pu", the on-off valve 103 cuts off the oil passage 102 and supplies the oil quantity Qs from the flow dividing valve 60 to the working machine hydraulic circuit 10 (non-unloaded condition). On the other hand, when the operating oil pressure Pw is defined by "Pmax.gtoreq.Pw&gt;Pu", the on-off valve 103 provides communication from the flow dividing valve 60 to the oil passage 102, and drains the oil quantity Qs into the tank 90 from the flow dividing valve 60 (unloaded condition). Irrespective of the magnitude of the operating oil pressure Pw, when a steering operation is effected at the minimum idling engine speed, the flow dividing valve 60 supplies the oil quantity Qs into the steering hydraulic circuit 20.
The operation of the aforesaid conventional art will be explained with reference to FIGS. 19A, 19B, 19C, and 20. In order to make the explanation easier, it is assumed that the oil flow Qs flows to the working machine hydraulic circuit 10 side and that the reduction ratio from the engine 70 to each of the hydraulic pumps 30, 40, and 50 is "1", unless otherwise specified.
In FIG. 19A, the operating oil pressure Pw is plotted in the axis of ordinates, and the displacement volume Vw of the working machine hydraulic pump 30 per one rotation of the engine 1 is plotted in the axis of abscissa. In FIG. 19B, the operating oil pressure Pw is plotted in the axis of ordinates, and the displacement volume Vs of the auxiliary hydraulic pump 50 per one rotation of the engine 1 is plotted in the axis of abscissa. In FIG. 19C, the steering oil pressure Pst is plotted in the axis of ordinates, and the displacement volume Vst of the steering hydraulic pump 40 per one rotation of the engine 1 is plotted in the axis of abscissa. The relief pressures of the circuits 10 and 20 are not necessarily the same, but in the present embodiment both of them have the same pressure Pmax.
FIG. 19A shows the oil pressure torque Tw (=Pw*Vw) of the working machine hydraulic pump 30 per one rotation of the engine 70. FIG. 19B shows the oil pressure torque Ts (=Pw*Vs) of the auxiliary hydraulic pump 50 per one rotation of the engine 70. The oil pressure torque Tws (not illustrated) is the total of the oil pressure torque Tw and the oil pressure torque Ts, i.e., "Tws=Tw+Ts". This can be also considered as "Tws=Pw*(Vw+Vs)" in a non-unloaded condition (Pw.ltoreq.Pu). The maximum value of the oil pressure torque Tws in a non-unloaded condition occurs at the time when "Pw=Pu" ("hatched portions shown by the diagonal lines extending upwardly to the right in FIGS. 19A and 19B). On the other hand, in an unloaded condition (Pmax.gtoreq.Pw&gt;Pu), the oil flow Qs discharged from the auxiliary hydraulic pump 50 is drained into the tank 90 by means of the unload circuit 100; therefore, "Ts=0" holds good. Accordingly, the oil pressure torque Tws is defined by "Tws=Tw", specifically, "Tws=Pw*Vw". The maximum value of the oil pressure torque Tws in an unloaded condition occurs at the time of "Pw=Pmax" ("hatched portions shown by the diagonal lines extending upwardly to the left in FIG. 19A). On the other hand, the oil pressure torque Tst of the steering hydraulic pump 40 per one rotation of the engine 70 is defined by "Tst=Pst*Vst".
In FIG. 20 the engine torque Te is plotted in the axis of ordinates while the engine speed Ne is plotted in the axis of abscissa. FIG. 20 shows the total amount of the oil pressure torque (Tw+Ts+Tst=Tws+Tst) relative to the engine torque Te, an area of excess torque .DELTA.T1 relative to the engine torque of the vehicle, and distribution examples of the excess torque .DELTA.T1 and traveling torque Ttr.
The disadvantages of the aforesaid conventional art are as follows.
(1) During operation with a high load, there is a disadvantage of reducing the operational efficiency due to waiting time occurring, since the working machine hydraulic circuit 10 does not receive the oil flow Qs from the flow dividing valve 60, although a little higher flow rate of oil would speed up the operation. The operation with a high load, e.g., excavating rock or the like with a bucket, corresponds to an unloaded condition (Pmax.gtoreq.Pw&gt;Pu).
(2) The actual operation of the vehicle is mainly an operation with a low load, such as raising the boom with a load in a bucket. The actual operation has an extremely high percentage of non-unloaded condition (Pw.ltoreq.Pu). Therefore, in order to increase the operational speed in the non-unloaded condition and further to reduce relief loss of the oil pressure at the time of "Pw=Pmax" in an unloaded condition, the vehicle is normally designed to have "(Pu*(Vw+Vs))&gt;(Pmax*Vw)". In order to efficiently use the engine torque Te, the volume of the bucket, the inner diameter of each of the hydraulic cylinders 11 and 12, and the like, of the vehicle are designed so that a condition close to "Tws=(Pu*(Vw+Vs))", (specifically, a condition close to "Pw=Pu") becomes normal. Further in the actual operation, there are a great many occasions in which the traveling and steering operations are carried out at the same time. Accordingly, as FIG. 20 shows, the excess torque .DELTA.T1 (=Te-(Tw+Ts+Tst+Ttr)) becomes smaller. In addition, as for the vehicle, switching operations between traveling forwardly and traveling rearwardly are frequently carried out in the normal condition. In this situation, an operator lets up on the accelerator pedal to reduce the engine speed Ne, and presses downwardly on the accelerator pedal to increase the engine speed Ne, directly after switching to traveling forwardly or traveling rearwardly. However, the excess torque .DELTA.T1 is small; therefore, the accelerating performance of the engine speed Ne is worse, and the resulting disadvantages include that the engine 70 produces black smoke, the fuel consumption efficiency becomes worse, and the like occur. The realities are that, since the excess torque .DELTA.T1 is secured in the limited engine torque Te, it cannot be made indiscriminately great.
(3) Further, in the aforesaid (2), when the operator lets up on the accelerator pedal to reduce the speed of the engine 70 to the minimum idling engine speed and a steering operation is conducted, the oil flow Qs, discharged from the auxiliary hydraulic pump 50, is switched and flows to the steering hydraulic circuit 20 via the flow dividing valve 60 instead of to the working machine hydraulic circuit 10. Accordingly, the speed of the working machine is reduced, but the steering speed becomes higher. In such an operation, it is preferable that the steering speed be low in order to prevent a load from falling from the bucket. However, as described above, the oil flow Qs is supplied to the steering hydraulic circuit 20 by the flow dividing valve 60 so that a sufficient quantity of oil is secured in the steering hydraulic circuit 20. For this reason, the operator tends to increase the steering speed, and a load tends to fall off the bucket. On the other hand, if the operator controls the steering speed to prevent the load from falling off, the oil discharge Qs from the auxiliary hydraulic pump 50 becomes excessive in the steering hydraulic circuit 20. The excess quantity of oil is diverted to the working machine hydraulic circuit 10 by way of an oil pressure torque lost portion, such as an orifice provided in the flow dividing valve 60. Specifically, a load is prevented from falling off, but oil pressure torque is lost. It is natural that the excess torque .DELTA.T1 is not increased.
There is a known art in which the engine power of a working vehicle is used for traveling by the medium of a torque converter, and is used for driving a working machine by the medium of a working machine pump. FIG. 21 shows an example of the engine torque characteristics of the working vehicle; the engine speed N is plotted in the axis of abscissa; and the engine torque T is plotted in the axis of ordinates. With an engine output torque Ta, the absorption torque (tractive torque) Tb of a torque converter becomes the tractive torque of the vehicle by the medium of the torque converter. In such a working vehicle, a working machine pump is generally a fixed displacement type. According to a first conventional art, when the load becomes heavier, as a result of an increase in the working machine torque Tc (proportional to the product of the load pressure P and the pump capacity V of the working machine pump), the acceleration torque Td, which is the result of subtracting the tractive torque Tb and the working machine torque Tc from the engine output torque Ta, is decreased. In contrast to the first conventional art, a second conventional art, which increases the acceleration performance at lower engine speeds, is known. According to the second conventional art, the working machine torque Tc1 is reduced more as the engine speed becomes lower (the reduction quantity is shown by the diagonal hatching lines in FIG. 21); and the acceleration torque Td1, which is the result of subtracting the tractive torque Tb1 and the working machine torque Tc1 from the engine output torque Ta, is increased.
According to the second conventional art, shown in FIG. 22, a working machine pump 202, driven by an engine 201, is connected to a tank by way of a bleed-off conduit 214 when an operation valve 203 is in its center valve position. When the operation valve 203 is in its operational position a or b, the working machine pump 202 is connected to a working machine actuator (for example, a hydraulic cylinder) 204a via supply conduits 215. A working machine pump capacity control means 205 has: a capacity control valve 205c, having a pilot pressure receiving portion 205f; a capacity control cylinder 205d; and a capacity control spring 205e coupling these components. The working machine load pressure P, outputted by the working machine hydraulic pump 202, is applied to the capacity control valve 205c and to a rod side of the capacity control cylinder 205d via an original pilot pressure conduit 221. In an engine speed detecting means 207, when the speed N of the engine 201 is reduced, the flow discharged from a fixed displacement pump 207a is reduced; therefore, the upstream pressure of an orifice 207b is reduced, and a pilot pressure control valve 207d is switched to its position b. For this reason, the pilot pressure, which is inputted to the pilot pressure receiving portion 205f of the capacity control valve 205c, is increased so as to be close to the pilot original pressure which is controlled at a fixed pressure by the relief valve 207c. Specifically, as the engine speed N is decreased, the working machine pump capacity VL is reduced.
However, the second conventional art has the following disadvantages.
(1) Irrespective of the magnitude of the load pressure P of the working machine pump 202, as the speed of the engine 201 is lowered, the working machine pump capacity VL (cc/rev) is decreased. For this reason, with a light load, as in raising a bucket without a load, there is a disadvantage in that the speed of the working machine is decreased and the working efficiency is reduced, although the working machine pump capacity V has a surplus by the degree that the load pressure P is reduced due to the working machine torque Tc1 being the same.
(2) When the center bypass bleed off type operational valve 203, illustrated in FIG. 22, is operated to its operational position a or b from its center valve position c and a supply conduit 215 to the hydraulic cylinder 204a is opened by closing the bleed-off conduit 214, the discharge Q (CM.sup.3 /min) from the working machine pump 202 is smaller at a lower engine speed. For this reason, it is difficult to increase the oil pressure of the supply conduit 215 to the hydraulic cylinder 204a, and there is a disadvantage of increasing a dead zone of an operational lever until the working machine starts to move, thereby worsening operability.